Dual-flow centrifugal pump

ABSTRACT

A double-flow centrifugal pump is disclosed, in which, using an impeller, a fluid can be suctioned from two axial sides from a negative pressure area and can be delivered in the radial direction into a positive pressure area. The negative pressure area is sealed off with respect to the positive pressure area by at least one stationary pump component, for example the pump housing and/or an insert part. According to the invention, sealing gaps are formed as axial gaps arranged axially between the impeller and the pump component and extending in the circumferential direction and in the radial direction, wherein the gap width(s) of said axial gaps is less than the radial spacing (a) of the impeller from all the components arranged with a radial spacing from the impeller.

FIELD OF THE INVENTION

The invention relates to a preferably single-stage, double-flowcentrifugal pump, in particular a cooling water pump for a marine dieselengine or a ballast water delivery pump on a ship, with a pump housingand with a dual-flow impeller rotationally fixedly arranged on arotationally driven shaft, with which impeller a fluid can be suctionedfrom two axial sides from a negative pressure area (suction side) anddelivered in the radial direction into a positive pressure area(pressure side), whereby the negative pressure area is sealed off withrespect to the positive pressure area by means of at least two sealinggaps which are spaced apart axially (via the positive pressure area) andwhich are formed between the impeller and at least one stationary pumpcomponent, in particular the pump housing.

BACKGROUND OF THE INVENTION

In known dual-flow centrifugal pumps the sealing gaps, formed as annulargaps, run in the axial direction and are formed between the impeller andthe pump housing. During the operation of the known centrifugal pumps aresulting radial force component acting on the shaft supported on oneside occurs, in particular if the centrifugal pumps are not operated attheir optimal working point, so that the shaft with the impellerrotationally fixed on it is deflected in the radial direction. In orderto prevent the impeller from contacting the pump housing during thisdeflection movement the sealing gaps formed as axial gaps must bedimensioned to be appropriately wide. However, this results in aperformance loss of the pump since constantly delivered medium from theradial positive pressure area flows in the axial direction through thesealing caps into the negative pressure area (suction area). As aconsequence, the efficiency of the known centrifugal pumps issignificantly impaired. Previously cited centrifugal pumps are onlysuited for applications, if the shaft is supported on one side, in whichcomparatively low volume flows are to be delivered. In the case ofdual-flow centrifugal pumps for large-volume flow applications, forexample, for example in the case of cooling water pumps for a marinediesel engine or of ballast water delivery pumps on a ship, the shaftcarrying the impeller is supported as a rule on both axial sides of theimpeller in order to minimize the radial deflection movement duringoperation. In the case of an only one-sided support of the shaft forthese applications a shaft would have to be used with an appropriatelylarge diameter and/or a complex support.

SUMMARY OF THE INVENTION

Starting from the previously cited prior art, the invention solves thebasic problem of indicating a dual-flow centrifugal pump, in particularfor large-volume flows of at least 500 m³/h with which a high degree ofefficiency is possible without complex constructive measures. Thecentrifugal pump shaft carrying the impeller should preferably besupported exclusively on one side and have the smallest possiblediameter. A striking of the impeller on the pump housing should bereliably avoided.

This problem is solved by a generic, dual-flow centrifugal pump in whichthe sealing gaps are formed as axial gaps extending in thecircumferential direction as well as in the radial direction andarranged axially between the pump component and the impeller, the gapwidth of which gaps, preferably measured in the axial direction, isgreater than the radial distances of the impeller to all components thatare arranged with radial spacing radially outside of the impeller. Inother words, the gap width of the sealing gaps formed as axial gaps isgreater than the gap widths of all other gaps (radial gaps) that arelimited on one side by the impeller. This means, expressed another way,that the distances, measured in the radial direction, between theimpeller and any components of the pump are greater than the gap widthof the sealing gaps formed as an axial gap.

Advantageous further developments of the invention are indicated in thevarious claims. The scope of the invention includes all combinations offeatures disclosed in the specification, the claims and/or in thefigures. In order to avoid repetitions, features disclosed as devicesare considered as disclosed in accordance with the method and can beclaimed as such. Likewise, features disclosed in accordance with themethod are considered as disclosed in accordance with the devices andcan be claimed as such.

The invention is based on the concept that the sealing gaps between theimpeller and at least one pump part with which the suction side of thecentrifugal pump is sealed relative to the pressure side are to beconstructed running in the radial direction relative to its longitudinalextent, i.e., as an axial gap. In other words, the impeller inaccordance with the invention is distanced in the axial direction by thesealing gaps from the at least one, preferably exclusively one pumpcomponent. The width of the sealing gap, which width extends at leastapproximately in the axial direction, is less at least at one position,preferably over its longitudinal extent, than the distance between theimpeller and all other pump components arranged with a radial distanceto the impeller. In other words, the gap width of the sealing gap isless than the radial distance of the impeller to all pump componentslocated radially outside of the impeller. The sealing gaps aredistinguished in that their axial extent is (considerably) less thantheir radial extent. The gap width of the axial gap, (sealing gap)measured in the axial direction, is greater than the gap width, measuredin the radial direction, of a radial gap arranged between the impellerand the pump component limiting the axial gap.

The gap width of the sealing gap is preferably at least 20%, morepreferably at least 12% and even more preferably 6% of the radialdistance of the impeller 7 to the pump component limiting the axial gap,in particular to the pump housing and/or to an insertion part preferablyforming a housing section.

It is, of course, possible to provide several sealing gaps formed as anaxial gap on both axial sides of the radial discharge area of theimpeller. However, it is preferred to provide only one sealing gapformed as an axial gap, whereby the gaps with the smallest gap width areunderstood to be sealing gaps.

A variant of an embodiment is quite especially preferred in which thepreferably exclusively two sealing gaps are arranged in an area radiallyinside circumferentially closed radial gaps via which the impeller isdistanced from the at least one, preferably exclusively one pumpcomponent. It is especially preferable if the axial gaps extend startingfrom the radial gaps in the radial direction inward. Therefore, avariant of an embodiment is especially preferred in which the axial gapshave, at least in a radially inner area, a lesser distance to the shaftthan the radial gaps do. The sealing gaps are advantageously locatedinside an imaginary circular cylinder whose generated surface receivesthe radial gaps. As a result of such a variant of an embodiment thesealing action is improved.

It is especially purposeful if the impeller has a circular, cylindricalcasing contour, whereby it is even more preferred if the sealing gaps(axial gaps) are formed between a front side of the impeller comprisinga cylindrical casing contour and between the at least one, preferablyexclusively one pump component.

Alternatively, a casing contour can also be provided in which theimpeller extends with its discharge area further to the outside in theradial direction. However, as will be explained later, it is alsopreferred in the case of such a geometry if the axial sealing gap isarranged in an area that has a lesser radius than a possible radial gaparranged between the pump jet and the impeller.

Based on the formation of the sealing gaps as axial gaps, it is possibleto dimension the gap width of the sealing gaps considerably smaller thanin the prior art without there being the danger that the impellerstrikes the pump component limiting the sealing gaps upon a radialdeflection. It is therefore possible by means of the design of thesealing gaps in accordance with the invention to achieve a highefficiency of the centrifugal pump since the amount of liquid that flowsfrom the pressure area into the suction area (negative pressure area) isminimized by the small gap width of the sealing gaps. The distancebetween the impeller and the pump component and/or other components ofthe pump can be dimensioned in the radial direction in such a mannerthat even in case of the greatest possible deflection of the impelleroccurring during operation there is no danger of collision. It istherefore possible, even for applications with a large-volume flow, inparticular for marine applications, to realize an only one-sided supportof the impeller shaft since greater radial deflections of the impellercan be accepted than previously. Furthermore, the dimensioning of theshaft as such can be minimized.

An embodiment has an especially simple construction and is thereforepreferred in which the sealing gaps run—in the framework of thetolerances—exactly in the radial direction relative to theirlongitudinal extent. However, a slightly curved or slightly obliquedesign of the sealing gaps by an appropriate shaping of at least onestructural component (impeller and/or pump component, especially pumphousing) limiting the sealing gaps is possible, in particular in such amanner that the gap geometry of the curved deflection movement of theimpeller takes place especially with a one-sided shaft support so thatthe gap width remains at least independently constant independently ofthe degree of the deflection of the impeller during operation. Theradius of curvature corresponds in an especially preferred manner atleast approximately to the distance of the impeller to the support ofthe shaft carrying the impeller.

It is advantageously provided in a further development of the inventionthat the gap width of the sealing gaps formed as axial gaps is selectedfrom a value range between 200 μm and 2000 μm, quite especiallypreferably between 200 μm and 400 μm.

It is especially advantageous if the minimal, i.e., the slightest radialdistance of the impeller to the pump component of the centrifugal pump,which pump component limits the sealing gaps formed as axial gaps, (withthe impeller standing still) is selected from a value range between 2 mmto 10 mm. In other words, the distance between the impeller and thepreviously cited pump component is preferably greater than the distancesof the indicated value range. The previously cited minimal radialdistance is especially preferably not only the minimal radial distanceof the impeller to the at least one, preferably exclusively one pumpcomponent limiting the sealing gaps, but rather the minimal radialdistance of the impeller to all structural components of the pump, inorder to reliably prevent a collision upon a radial deflection.

An embodiment of the dual-flow centrifugal pump has an especiallypreferred construction in which the sealing gaps are arranged betweenthe front sides of the impeller, which front sides face in the axialdirection, and between the at least one pump component. In other words,it is preferred if the sealing gaps have the greatest possible axialdistance from each other. This can be realized, for example, in that theimpeller has a casing contour that is at least approximately circularlycylindrical. An imaginary generated surface of a centrifugal cylinderwhich surface receives the radial gaps especially preferably surroundsthe axial gaps radially on the outside.

As initially explained, the phrase an axial gap (sealing gap) extendingin the radial direction denotes not only an embodiment in which thesealing gaps run exactly in the radial direction relative to thelongitudinal extent, in the framework of the tolerances, that is, theyare constructed, for example, in the shape of annular disks. Anembodiment is also conceivable in which the sealing gaps have a slightangle of rise or are slightly curved, i.e., have a large radius ofcurvature that preferably corresponds at least approximately, inparticular in the case of a shaft supported on one side, to the distanceof the particular sealing gap from the shaft support. The particularsealing gap is thus designed in such a manner that the gap width duringthe operation of the centrifugal pump does not change or changes only asslightly as possible, that is, upon a possible radial deflection of theimpeller since the gap geometry follows the deflection movement. Thecurvature or beveling of the sealing gap can be realized by anappropriate geometric shaping of the impeller and/or of the at leastone, preferably exclusively one pump component limiting the sealing gapon the axial side opposite the impeller. The angle (angle ofinclination) of the particular sealing gap to an imaginary radial planearranged orthogonally to the longitudinal extension of the shaft isquite especially preferably selected from a value range between 0.01°and 2.0°. A possible radius of curvature is preferably selected from avalue range between 200 mm and 1000 mm, preferably 300 mm and 700 mm.

The radius of curvature of the particular sealing gap, more precisely atleast of a surface (of the impeller and/or of the pump component)limiting the sealing gap preferably corresponds at least approximatelyto the distance of the particular sealing gap (in particular on aradially innermost area of the sealing gap) to the shaft support, inparticular in the case of one (pump shaft) supported on one side. In acorresponding manner the angle of inclination of the gap explained inthe specification refers to the angle of at least one surface (of theimpeller and/or of the pump component) limiting the sealing gap relativeto the previously cited radial plane.

As already indicated, it is an extremely cost-effective variant of anembodiment of the centrifugal pump if the shaft carrying the impeller issupported exclusively on one side, preferably on an upper side.

It is especially purposeful if the centrifugal pump in accordance withthe invention is designed for large-volume flow applications, inparticular marine applications. The centrifugal pump is preferablydesigned for delivering a volume flow from a value range betweenapproximately 500 m³/h and approximately 4000 m³/h, preferably betweenapproximately 800 m³/h and approximately 1500 m³/h (for example, in thecase of rather small cooling water pumps) or between approximately 1500m³/h and approximately 2300 m³/h (for example, in the case ofaverage-size cooling water pumps) or between 2300 m³/h and 3500 m³/h(for example, in the case of rather large cooling water pumps),preferably at a maximum delivery level from a value range betweenapproximately 20 m and approximately 50 m, preferably from approximately30 m. It is especially preferable for reasons of space, especially formarine applications, if the dual-flow centrifugal pump is realized in avertical construction, that is, in such a manner that the shaft runsvertically to a standing surface of the centrifugal pump.

It is especially preferable if the centrifugal pump is a single-stagecentrifugal pump, that is, a pump comprising exclusively one impeller.

It is especially purposeful if the pump housing is a so-called spiralhousing that sets the flow path on the suction side to the two axialsides of the impeller and combines preferably two outlet conduits in ahelical manner on the pressure side.

The invention also comprises the use of a dual-flow centrifugal pumpdesigned in accordance with the concept of the invention as a coolingwater pump for a marine diesel engine or as a ballast water deliverypump on a ship.

Other advantages, features and details of the invention result from thefollowing specification of preferred exemplary embodiments as well asfrom the figures.

BRIEF DESCRIPTION OF THE DRAWINGS

The accompanying drawings illustrate preferred embodiments of thedisclosed method so far devised for the practical application of theprinciples thereof, and in which:

FIG. 1 shows a cross-section view of an exemplary embodiment of adual-flow centrifugal pump constructed in accordance with the concept ofthe invention,

FIG. 2 shows a schematic illustration of gap conditions for the pump ofFIG. 1, and

FIGS. 3-7 show a plurality of embodiments of the sealing gaps.

In the figures the same elements and elements with the same function arecharacterized with the same reference numerals.

DETAILED DESCRIPTION

FIG. 1 shows a sectional view of a dual-flow centrifugal pump 1 in avertical orientation (i.e., the impeller shaft is oriented vertically).The exemplary embodiment shown is a cooling water pump for a marinediesel engine designed for delivering a volume flow of 2300 m³/h at amaximum delivery level of 30 m.

The centrifugal pump 1 comprises a pump housing 2 designed as a spiralhousing and with a suction-side inlet 3 as well as a pressure-sideoutlet 4. A shaft 5 supported on one side extends into the pump housing2 from above downward in a vertical direction and is supported by abearing 6 constructed as a ball bearing. The shaft 5 carries on itsinside a dual-flow impeller 7 with a substantially circularlycylindrical casing contour. The impeller 7 sits in a rotationally fixedmanner on the shaft 5. A shaft seal 8 is located in an area axiallybetween the support 6 and the impeller 7. As is apparent from FIG. 1 theshaft 5 extends in an area above the shaft seal 8 through a cover 9fixed by screwing on the pump housing 2.

The impeller 7 separates a negative pressure area 10 (suction side) froma positive pressure area 11 (pressure side).

The shaft 5 can be rotated in a known manner by an engine (not shown),in particular by an electromotor, whereby the impeller 7 rotating withthe shaft 5 sucks fluid, here cooling water, from both axial sides outof the negative pressure area 10 and delivers it in a radial directionoutward into the positive pressure area 11, whereby the positivepressure area 11 is subdivided into two helically arranged flow conduits12, 13 separated from one another by a dividing wall 14. The two flowconduits 12, 13 and the fluid currents are brought together again in thearea of the outlet 4.

During the operation of the centrifugal pump 1, especially when thecentrifugal pump 1 is not working at an optimal working point, a loadingof the shaft 5 with a radial force occurs in the area of the impeller 7which load has the tendency to deflect the shaft 5 with impeller 7 inthe radial direction. In order to prevent the impeller 7 from collidingwith the pump housing 2 (pump component) in the radial direction twoaxially spaced radial gaps 15, 16 extending in the axial direction aredimensioned to be so wide that even a maximally conceivable deflectionof the shaft 5 during operation cannot result in a collision of theimpeller 7 with the pump housing 2. The radial gaps 15, 16 are notdesigned as sealing gaps and fulfill no sufficient sealing function onaccount of their comparatively large gap width (measured at thenarrowest position) from the one in the exemplary embodiment shown ofapproximately 5 mm. The radial gaps have the form of circular,cylindrical surfaces. If the radial gaps 15, 16 were the only sealinggaps, the centrifugal pump 1 would have an extremely poor efficiency onaccount of the comparatively large gap width since liquid, here coolingwater, would constantly flow in a large amount through the radial gaps15, 16 from the positive pressure area 11 into the negative pressurearea 10 and would thus be directly delivered in a circuit.

In order to achieve the desired sealing action while avoiding the dangerof a collision between impeller 7 and pump housing 2 (pump component)the pump housing 2 (pump component) extends over the impeller 7 on bothaxial sides, i.e., above and below in an inward radial direction in sucha manner that a sealing gap, 19, 20 that is constructed as an axial gapand extends as regards its longitudinal extension in the radialdirection is formed between each front side 17, 18 of the impeller 7 andthe pump housing 2 (pump component). It is essential that these sealinggaps, 19, 20, measured at their narrowest position, have a smaller gapwidth than the radial gaps 15, 16.

The sealing gaps, 19, 20 are located radially inside the radial gaps 15,16, whereby the radial gaps 15, 16 merge into the sealing gaps, 19, 20and the sealing gaps, 19, 20 border directly on the radial gaps 15, 16.In the exemplary embodiment shown the width of the sealing gaps 19, 20is about 400 μm.

The sealing gaps, 19, 20 are, as explained, limited on the one hand inthe axial direction by the impeller 7, in the exemplary embodiment shownby a front side 17, 18 of the impeller 7 and on the opposing side by awall surface 21, 22 of the pump housing 2, which wall surface is alignedhere parallel to the particular front side 17, 18.

If a deflection of the impeller 7 occurs in a radial direction duringoperation the front sides 17, 18 are shifted substantially parallel tothe wall surfaces 21, 22 of the pump housing 2, so that no collision canoccur there. The radial gaps 15, 16 are, as explained, dimensioned to beso wide that even here a collision with the impeller 7, even at amaximally admissible deflection, is excluded.

FIG. 2 schematically shows the gap conditions.

The schematically shown impeller 2 can be recognized, arranged in arotationally fixed manner on a rotatably supported shaft 5.

The impeller 7 is surrounded by a pump component 23, here the pumphousing 2, more precisely an inserted part 24 that forms a component ofthe pump housing 2. Alternatively, the inserted part 24 may not beconstructed and arranged to form a component of the housing, therefore,inside the pump housing and at a distance to an outer housing side. Upona rotation of the impeller 7 the liquid flows in the arrow directionsfrom the suction side (negative pressure area) 10 to the pressure side(positive pressure area) 11.

Two sealing gaps 19, 20 are formed between the pump component 23, thatcan be designed to be monopartite or bipartite, and the impeller 7, moreprecisely between the front sides 17, 18 of the impeller 7 comprising acircular, cylindrical casing contour. These sealing gaps 19, 20 areaxial gaps that are formed axially between the pump component 23 and theimpeller 7. The gap widths “s” of the sealing gaps 19, 20 are 400 μm inthe exemplary embodiment shown. The two sealing gaps 19, 20, in the formof flat annular disks are distanced from one another in the axialdirection and, among other things, separated from one another by theradial outlet area or areas of the impeller 7.

In the exemplary embodiment shown, in addition to the sealing gaps 19,between the impeller 7 and the pump component 23 two radial gaps 15, 16are provided whose gap width a is greater than the gap width s of thesealing gaps. In the exemplary embodiment shown the gap widths “a” areapproximately 5 mm when impeller 7 is at a standstill. The sealing gaps19, 20 are located radially inside the radial gaps 15, 16 and thereforeare closer to the shaft 5 (i.e., smaller) than the radial gaps 15, 16.The radial gaps have the form of circular, cylindrical jacket. Thesealing gaps 19, 20 have approximately the form of a circular ring disk.Providing the (narrow) radial gaps 15, 16 can also be eliminated in amodified constructive designed of the pump component 23. It is alsoconceivable to provide several sealing gaps 19, 20 present in parallelplanes that are axial gaps on at least one of the two axial sides,preferably on both axial sides of the impeller 7. In such cases, twoaxially adjacent sealing gaps are preferably connected to one anothervia a radial gap with a larger gap width than the gap width of thesealing gaps on at least one axial side of the impeller 7. Thus, astepped gap formation would result, whereby the axial gap section wouldrepresent the sealing gaps. Therefore, a stepped gap design results.

Conceivable alternative sealing gap geometries are shown in the FIGS. 3to 7, whereby the angles or radii of curvature are shown in anexaggerated manner for reasons of clarity. In reality, minimal rises andlarge radii of curvature are involved.

All exemplary embodiments have the fact in common that the sealing gapsare axial gaps that run (as regards their longitudinal extension)substantially in the radial direction and their axial extension is(substantially) less than their radial extension.

In the exemplary embodiment according to FIG. 3 a sealing gap 19 isformed between the impeller 7 and a pump component 23. The section ofthe impeller 7 limiting the sealing gap 19 runs relative to thelongitudinal extension of the shaft exactly in the radial direction,whereas on the contrary the surface section of the pump component 23,that limits the sealing gap 19, is slightly inclined relative to aradial plane, here under an angle α of <1°. This results in a sealinggap inclination about this angle α relative to an imaginary radial planein which in the exemplary embodiment shown the represented surfacesection of the impeller 7 lies.

In the exemplary embodiment according to FIG. 4 the surface section ofthe impeller 7 limiting the sealing gap 19 as well as the surfacesection of the pump component 23 that is opposite and limits the sealinggap 19 are inclined relative to a radial plane, in the exemplaryembodiment shown both under the same angle α of <10° here. The use ofdifferent but similar angles of inclination is also possible.

In the exemplary embodiment according to FIG. 5, the surface area of theimpeller 7 limiting the sealing gap 19 is located in a radial planerelative to the longitudinal extension of the shaft, in contrast towhich the surface area of the pump component 23 limiting the sealing gap19 is curved, which curvature preferably has a radius that has thesealing gap 19 from the support of the shaft 5 (not shown).

In the exemplary embodiment according to FIG. 6 both surfaces limitingthe sealing gap 19 as well as the surface of the impeller 7 and thesurface of the pump component 23 are designed to be slightly curved.

In the exemplary embodiment according to FIG. 7 the surface of theimpeller 7 limiting the sealing gap 19 is designed level but inclined atan angle α of <10° to the radial plane, in contrast to which the surfaceof the pump component 23 limiting the sealing gap 19 is slightly curvedand preferably has a radius of curvature of 500 mm.

1. A double-flow centrifugal pump for delivering a flow rate in a rangefrom 500 m³/h to 4000 m³/h, the pump positioned in a verticalorientation, and having a pump housing and a dual-flow impellerrotationally fixedly arranged on a rotationally driven shaft, with whichimpeller a fluid can be suctioned drawn from two axial sides from anegative pressure area and delivered in the radial direction into apositive pressure area, wherein the negative pressure area is sealed offwith respect to the positive pressure area by at least two axiallyspaced sealing gaps formed between the impeller and at least onestationary pump component, the at least one stationary pump componentcomprising at least one of the pump housing and a stationary insert, andwherein the shaft is mounted only on one side, and the sealing gapscomprise as axial gaps extending in the circumferential direction aswell as in the radial direction and arranged axially between theimpeller and the at least one stationary pump component, the gap widthof which is less than the radial distance of the impeller to allcomponents that are arranged at a radial distance to the impeller. 2.The centrifugal pump according to claim 1, wherein the gap width of thesealing gap is between 200 μm and 2000 μm.
 3. The centrifugal pumpaccording to claim 1, wherein the radial distance of the impeller to theat least one stationary pump component of the centrifugal pump with theimpeller standing still is between 2 mm to 10 mm.
 4. The centrifugalpump according to claim 1, wherein the sealing gaps are arranged betweenthe front sides of the impeller and between the pump housing or thestationary insert.
 5. The centrifugal pump according to claim 1, whereinthe sealing gaps run exactly in the radial direction.
 6. (canceled) 7.The centrifugal pump according to claim 1, wherein the centrifugal pumpis configured to deliver a volume flow in a range between approximately800 m³/h and approximately 1500 m³/h.
 8. (canceled)
 9. The centrifugalpump according to claim 1, wherein the two sealing gaps have a smallerradial distance from the shaft than radial gaps between the impeller andthe at least one stationary pump component.
 10. The centrifugal pumpaccording to claim 1, wherein the impeller has a circular, cylindricalcasing contour.
 11. The centrifugal pump according to claim 1, whereinthe centrifugal pump is configured to deliver a volume flow in a rangebetween about 1500 m³/h and about 2300 m³/h at a maximum delivery level.12. The centrifugal pump according to claim 1, wherein the centrifugalpump is configured to deliver a volume flow in a range betweenapproximately 2300 m³/h and 3500 m³/h at a maximum delivery level. 13.The centrifugal pump according to claim 1, wherein the centrifugal pumpis configured to deliver said volume flow at a maximum delivery level ina range between about 20 m and about 50 m.
 14. The centrifugal pumpaccording to claim 13, wherein the maximum delivery level isapproximately 30 m.
 15. The centrifugal pump according to claim 1,wherein the sealing gaps include a radial plane an angle ofapproximately 0° and 1°.
 16. The centrifugal pump according to claim 1,wherein the sealing gaps have a radius of curvature of about 200 mm to1000 mm.
 17. The centrifugal pump according to claim 1, wherein thesealing gaps have a radius of curvature of about 300 mm to 700 mm.